Typically, the crankshaft of an automobile is driven (or rotated) by the reciprocating motion of the pistons of an internal combustion engine. The rotational motion of the crankshaft is then transferred to the wheels to drive the automobile. The combustion forces that are induced upon the crankshaft, and transferred through the piston-rod connection, introduce torque pulses that act to spin the crankshaft. Oftentimes the torque-pulse-excitation of the crankshaft occurs at a rate (or frequency) that corresponds with the crankshaft's first or second natural torsional frequency. If a crankshaft is left to operate in a high amplitude torsional resonance condition, the crankshaft is likely to fail much sooner than desirable. Therefore, it is typical to control a crankshaft that is operating in a resonant condition by adding a specially designed crankshaft damper. Typical crankshaft dampers include an inertial mass that is coupled to the crankshaft by an elastic element that possesses a torsional spring rate. The torsional spring rate of the elastic element is generally governed by the shear modulus of the elastic element.
In addition to torsional vibrations, dampers may also be used to manage bending vibrations that occur at the crankshaft's nose, also referred to here as the crank-nose. The bending amplitude of the crank-nose can be characterized as a vector summation of axial thrust (along the crankshaft's rotational axis) and planar loading (perpendicular to the crankshaft's rotational axis). The forcing frequency of the bending vibrations is typically firing-order driven and the bending amplitudes increase with increased engine RPM, increased cylinder pressure and/or harmonic resonance of the crank-nose. The effects of crank-nose bending can vary from no deleterious affects, to compromised front sealing, to main bearing wear, to crank-nose bending and breaking.
Although the crankshaft may experience both torsional and bending vibrations, the use of two dampers, including a first damper to control torsional vibrations and a second damper to control bending vibrations, is neither efficient, nor cost-effective. Specifically, the addition of a second inertial mass to control bending vibrations drains the engine's fuel economy and torque responsiveness away from its primary function as a power source. In addition, the use of more than one damper may increase costs. Accordingly, the use of a single inertia damper is desirable. Specifically, when a crankshaft damper is already in place to combat torsional vibrations, it is desirable to manipulatively design the existing damper, which has its own natural bending frequency, to offset the effects of crankshaft bending amplitudes as well.
One commonly recognized limitation in the design of a single inertia damper is the inherent interaction between torsional and bending damper frequency. Because the spring rate/shear rate of the elastic element in the torsional direction may govern the torsional frequency of the damper, and because the spring rate/shear rate of the elastic element in the axial direction may govern the bending frequency of the damper, if the spring rate/shear rate of the elastic element is the same in each direction, as is the case with most materials, then the torsional frequency of the damper cannot be altered independently of the bending frequency of the damper by simple material substitution. It is frequently the case, however, that the torsional frequency of the damper must be adjusted without altering the bending frequency, or vice versa. Accordingly, one single inertia damper design, which is disclosed in U.S. Pat. No. 5,231,893 (“the '893 patent”), functions by altering the geometric shape of the crankshaft-to-inertial mass joint to take advantage of the difference between the spring rate of the elastic element in shear and the spring rate of the elastic element in compression. Specifically, by altering the joint geometry the torsional spring rate may be governed by the shear rate of the elastic element while the bending spring rate may be governed by the modulus of compression of the elastic element. In this manner, altering the geometry of the damper permits independent adjustment of the bending and torsional damper frequencies.
A fundamental limitation of the single inertia damper design disclosed in the '893 patent is that the joint geometry must be redesigned in order to adjust the torsional damper frequency independently of the bending damper frequency. The need for such adjustment is not an exaggerated concern since torsional damper frequency changes are common during the course of engine development and sometimes even during mature-engine performance and use proliferation. The redesign of the joint geometry requires the additional cost and lead-time to re-tool and revalidate. Moreover, the torsional and bending frequency ranges that can be controlled by current geometry-altering dampers is limited by the range of geometric shapes that are available.
Accordingly, a new single inertia damper design is desired.